Electronic control device for a refrigerant compressor

ABSTRACT

The invention relates to an electronic control device ( 13 ) for a refrigerant compressor, comprising at least: a drive unit ( 18 ); and a compression mechanism ( 5 ) which is actively connected to the drive unit ( 18 ), with at least one piston ( 9 ) which is driven by a crankshaft ( 6 ) and moves back and forth between a lower and an upper dead point in a cylinder of a cylinder block ( 8 ), in which the electronic control device ( 13 ) is designed to detect, control and/or regulate the rotational speed (ω) of the drive unit ( 18 ) and to at least approximately detect the piston position, and in which the electronic control device ( 13 ) is designed to drive the compression mechanism ( 5 ) by means of the drive unit ( 18 ) in such a way that at least one drive angle segment (ΔΦ) and at least one transit angle segment (Δτ) is provided for the duration of a regulating time interval (Δt) comprising more than one crankshaft rotation, for a plurality of crankshaft rotations, preferably for each crankshaft rotation of the regulating time interval (Δt), and the compression mechanism ( 5 ) is subject to a positive operating torque (Bm) during the at least one drive angle segment (ΔΦ), and to a smaller positive operating torque (Bmv) compared to the positive operating torque (Bm) or to no positive operating torque (Bm) during the at least one transit angle segment (Δτ).

FIELD OF THE INVENTION

The present invention relates to an electronic control device for a refrigerant compressor, comprising at least:

a drive unit,

a compression mechanism that is actively connected to the drive unit, with at least one piston that is driven by a crankshaft and moves back and forth between a lower and an upper dead point in a cylinder of a cylinder block,

wherein the electronic control device is designed

to detect and to control and/or regulate the rotation speed of the drive unit, and

to detect the piston position.

The present invention also relates to a method for regulating a reciprocating piston refrigerant compressor, whose compression mechanism is driven by means of a drive unit with an operating torque, in particular, a method for regulating the drive unit of the reciprocating piston refrigerant compressor.

PRIOR ART

Such electronic control devices are used in refrigerant compressors with variable rotation speed, in particular, also in refrigerant compressors that build on the reciprocating piston principle. Refrigerant compressors with variable rotation speed have the advantage that they can be tuned specifically to the refrigeration requirements of the object to be cooled, in that they can be operated, for example, in the case of lower refrigeration requirements, at a lower rotation speed and, in the case of increased refrigeration requirements, at a correspondingly increased rotational speed.

The construction of reciprocating piston refrigerant compressors with variable rotational speed has been known for a long time. They consist essentially of a drive unit and a compression mechanism in the form of a piston that moves back and forth between an upper and lower dead point in a cylinder housing and that is connected by means of a connecting rod to a crankshaft that is, in turn, coupled rotationally fixed to a rotor of the drive unit.

Typically, a brushless direct-current motor is used as the drive unit. Here, it is possible to determine the relative position of the rotor of the direct-current motor and thus also the rotational speed of the motor or the compression mechanism on the basis of the countervoltage (induction countervoltage) induced in the motor winding. This method eliminates the need for separate sensors and is therefore especially easy to implement and less susceptible to interference.

In reciprocating piston refrigerant compressors, the normal regulated operation at low rotation speeds and the stopping process are very often problematic in the design.

The problems in the operation at low rotational speeds in normal regulated operation are also, among other things, noise-related.

In the intake and compression phases, different gas forces (caused by the refrigerant pressure relationships in the system) and friction forces (both together are designated as load torque) act on the compression mechanism, which, taking a closer look, results in a rotational frequency that is non-uniform, because it varies as a function of the crank angle. In the present application, there is a difference, in principle, between the terms “rotational frequency” and “rotational speed.” The term “rotational frequency” is used when the actual, instantaneous rotational frequency is meant, which varies in reciprocating-piston refrigerant compressors according to the state of the art with respect to the crank angle, while the term “rotational speed” is used when the average rotational frequency of one crankshaft rotation is meant, that is, the value that is generally meant when one speaks of the rotational frequency of a reciprocating piston refrigerant compressor.

In actuality, during the compression phase, a load torque that is increased relative to the intake phase acts on the compression mechanism, which must be overcome by the operating torque of the drive unit in order to keep the compression process in operation. The increased load torque in the compression phase has the result that, in reciprocating piston refrigerant compressors according to the state of the art, which are operated at constant voltage, the rotational frequency of the compression mechanism is reduced in the compression phase.

In contrast, during the intake phase, the gas forces cause a reduced load torque relative to the compression phase. This leads to an increase in the rotational frequency of the compression mechanism during the intake phase.

Thus, overall, a load torque varying with respect to the crank angle acts on the compression mechanism, wherein the fluctuation range of the load torque depends primarily on the compression ratio in the refrigerant circuit and leads to angular accelerations of different magnitudes and thus to a rotational frequency of the compression mechanism that is non-uniform with respect to the crank angle during one crankshaft rotation.

To compensate for oscillations and vibrations of the compression mechanism during the operation, this is supported in a housing, together with the drive unit, by means of spring elements. The natural frequencies of this oscillation system are between 5 Hz and 16 Hz, depending on the compressor type.

Thus, the increased load torque recurring during each crankshaft rotation during the compression phase, especially during the operation of the reciprocating piston refrigerant compressor at rotational speeds below a range between 1000 rpm and 700 rpm, leads to shocks on the compression mechanism, which compresses the compression mechanism together with the drive unit into the spring element and deflects this spring element, wherein the frequency of shocks is in the area of the natural frequency of the oscillation system, so that the deflections of the spring element increase with each crankshaft rotation such that the compression mechanism and/or the drive unit might impact against the housing, which can lead to undesired sound emissions. These circumstances are also a reason why known reciprocating piston refrigerant compressors are not operated below . range between 1000 rpm and 700 rpm in the normal regulated operating phase.

The described, undesired sound emissions of a reciprocating piston refrigerant compressor at low rotational speeds occur not only in the normal regulated operation, but primarily also during the stopping process, wherein the system must pass through these low rotational speeds. The stopping process usually runs as follows:

If, after an appropriate duration of a normal regulated operating phase of the refrigerant compressor, the target temperature of the object to be cooled, for example, a cooling compartment of a refrigerator, has been reached, the electronic control device of the refrigerator sends a signal to the electronic control device of the refrigerant compressor, with which it is reported to this control device that cooling power is no longer needed, because the target temperature has been reached. From the prior art it is known that the electronic control device of the refrigerant compressor then switches off the drive (switch-off time) and begins the stopping process.

The crankshaft of the compression mechanism runs through complete rotations even after the switch-off time beginning at the upper dead point (crank angle 0°), wherein initially an intake phase (correctly: intake and re-expansion phase) is performed, during which refrigerant is suctioned into the cylinder. The suction phase ends, theoretically when the cylinder has reached the bottom dead point (crank angle 180°). Then, the compression phase begins (correctly: compression and ejection phase, during which the refrigerant located in the cylinder is compressed and pushed out of the cylinder. The compression phase ends theoretically when the piston has reached the upper dead point (crank angle 360°) again. In practice, the actual compression of the refrigerant, however, begins only at a crank angle of approximately 210° (depending on the refrigerant compressor, the pressure ratios, the valve design, etc.), but in any event after 180°, and the intake phase begins at approximately 30°, but in any event after the upper dead point.

Switching off the drive unit of the refrigerant compressor at a switch-off time initiates the stopping process and has the result that the compression mechanism is in a driveless state (without operating torque) and continues to rotate only due to its mass inertia until it has come to a complete standstill, i.e., its rotational speed is 0. Colloquially, one could also say that the refrigerant compressor “runs out.”

During the driveless state, the compression mechanism/drive unit rotates only due to it kinetic energy and its mass inertia at the switch-off time. Thus, it rotates, so to speak, in an uncontrolled manner, and its rotational speed behavior is dependent on the load torque acting on the compression mechanism. The load torque leads to a reduction of the rotational speed of the driveless, running-out refrigerant compressor, so that the kinetic energy of the compression mechanism decreases and decreases until it is no longer sufficient, depending on the pressure ratios in the refrigerant circuit, to overcome the load torque (limit rotational speed).

Here, it is to be taken into account that when the drive unit is switched off, differently than during the normal regulated operating phase, there is no positive operating torque that counteracts the load torque, in particular, the increased load torque in the compression phase, so that when the drive unit is switched off, the shocks acting on the compression mechanism due to the increased load torque in the compression phase pass through, so to speak, unbraked, and therefore the effects are even more serious with respect to the deflection of the spring elements than is the case in the normal regulated operating phase, where the positive operating torque counteracts the shock and thus somewhat damps the shocks.

This, in turn, has the effect that the deflection of the spring element at low rotational frequencies during the stopping process is even larger that during the normal regulated operation of the reciprocating piston refrigerant compressor for the same low rotational frequencies, so the likelihood of contact between the compression mechanism/drive unit and housing is also higher, which is associated, overall, with higher noise emission.

In addition, for the case that the piston is located in a compression phase, the situation can occur that the kinetic energy of the compression mechanism/drive unit is no longer sufficient to overcome the load torque, and the piston of the compression mechanism is pressed back in the direction of the lower dead point, thus the direction of rotation of the compression mechanism is reversed.

The reversal of the direction of rotation is associated with an additional stopping jolt that acts on the compression mechanism and presses the compression mechanism/drive unit into the spring element and also deflects this element.

Just during the stopping process, where, as already described above, a positive operating torque does not counteract the load torque and the increased load torque acting in the compression phase in any event already excites the oscillation system in a jolt-like way in the region of its natural frequency, the stopping jolt due to the reversal of the direction of rotation contributes even more to the deflection of the spring element, so that the likelihood that the compression mechanism/drive unit will contact the housing wall is increased even more, and thus undesired noise emissions are produced.

From the prior art it is known to end the driveless phase by applying a braking torque and in this way to avoid at least a backward jolt of the refrigerant compressor and thus the stopping jolt. Specifically, from EP 2 669 519 A1 and DE 20 2012 013 046 derived therefrom, it is known to use a braking torque to brake the driveless rotating compression mechanism/drive unit after the switch-off time when the rotational speed falls below a certain rotational speed. For this purpose, it is necessary to constantly monitor the rotational speed of the driveless, rotating compression mechanism after the switch-off time and at a defined rotational speed that, in any event, still must be high enough to overcome the load torque, that is, must be above the limit rotational speed, to actively brake the compression mechanism by means of a braking torque that is applied to the compression mechanism.

To avoid the problems described above in connection with the jolts and the jolt frequency due to the increased load torque during the compression phase, the braking torque is applied already at rather high rotational speeds, which is energetically disadvantageous and also causes additional sound emissions.

Thus, from the mentioned reasons, it can be stated in summary that the operation of known reciprocating piston refrigerant compressors with variable rotational speeds at low rotational frequencies causes an excitation of the oscillation system in the range of its natural frequencies and thus leads to undesired sound emissions, wherein this effect is to be observed especially strongly during the stopping process during the driveless run-out period.

In connection with the stopping process of a reciprocating piston refrigerant compressor, the load torque that varies greatly depending on the crank angle produces the risk that, starting from a limit rotational speed, the compression stroke can no longer be completed and the compression mechanism strikes back in the opposite direction, whereby the oscillation system is also excited and the spring elements are also deflected, which causes, in turn, an increased stop likelihood of the compression mechanism/drive unit.

Task of the Invention

Therefore, the goal of the invention is to provide an electronic control device for a reciprocating piston refrigerant compressor and a method for regulating a reciprocating piston refrigerant compressor that enables the operation of a reciprocating piston refrigerant compressor even at low rotational speeds without causing excessive deflection of the spring element and without resulting in increased sound emissions.

Another goal of the invention is to provide an electronic control device for a reciprocating piston refrigerant compressor and also a method for regulating a reciprocating piston refrigerant compressor that enables stopping of the reciprocating piston refrigerant compressor that has been optimized with respect to noise without a braking torque having to be applied to the compression mechanism.

Another goal of the invention is to provide an electronic control device for a reciprocating piston refrigerant compressor and a method for regulating a reciprocating piston refrigerant compressor that enables the operation of a reciprocating piston refrigerant compressor even at low rotational speeds, in particular, when passing through such low rotational speeds during the stopping process, without the risk of reversing the direction of rotation and associated stopping jolts.

Another goal of the invention is to be able to tune the spring elements, by means of which the refrigerant compressor is supported in the housing, more specifically to the oscillations of the compression mechanism.

An additional goal of the invention is to provide an electronic control device for a reciprocating piston refrigerant compressor and a method of regulating a reciprocating piston refrigerant compressor that enables the operation of a reciprocating piston refrigerant compressor at a constant rotational frequency in order to damp the shocks acting on the compression mechanism due to the increased load torque during the compression phase, wherein the rotation speed can be simultaneously reduced or increased.

Another goal of the invention is to provide an electronic control device for a reciprocating piston refrigerant compressor and a method for regulating a reciprocating piston refrigerant compressor that enables the rotation speed of the compression mechanism to be decreased successively to a low rotational frequency during the stopping process without the possibly still occurring shocks, during the compression phase after the switching off in the driveless phase, causing a deflection of the spring element that is greater than the maximum permissible deflection at which a contact of the compression mechanism/drive unit on the housing is just avoided.

Presentation of the Invention

According to the invention, these goals are achieved in an electronic control device for a reciprocating piston refrigerant compressor that comprises at least a drive unit, a compression mechanism that is actively connected to the drive unit, with at least one piston that is driven by a crankshaft and moves back and forth between a lower and an upper dead point in a cylinder of a cylinder block, in which the electronic control device is designed to detect, control, and/or regulate the rotational speed of the drive unit and to detect the piston position, in that the electronic control device is designed to drive the compression mechanism by means of the drive unit in such a way that at least one drive angle segment and at least one transit angle segment are provided for the duration of a regulating time interval comprising more than one crankshaft rotation, for a plurality of crankshaft rotations, preferably for each crankshaft rotation of the regulating time interval (Δt), and the compression mechanism is subjected to a positive operating torque during the at least one drive angle segment and to a smaller positive operating torque compared to she positive operating torque or to no positive operating torque during the at least one transit angle segment.

The positive and/or reduced positive operating torque can either remain constant or vary during the drive angle segment and/or transit angle segment, wherein, in practice, a variation of the positive operating torque or reduced positive operating torque already occurs within tight limits due to the widespread control by means of constant voltage.

The control device according to the invention enables the compression mechanism to be controlled such that the load torque varying with respect to the crank angle can be at least partially equalized by a similarly varying, positive operating torque, wherein the electronic control device is designed such that the positive operating torque varies at least once with respect to a rotation of the crankshaft and beyond a regulating time interval that comprised several, preferably successive, preferably 5 to 15 successive crankshaft rotations. The positive operating torque can vary between a maximum value that is essentially dependent on the drive unit and the value zero.

In this way, it is possible to tune the drive angle intervals to the load torque so that a possibly uniform rotational frequency profile can be achieved for each crankshaft rotation. For example, the tuning can take place so that the at least one drive angle segment is provided within each operating phase of the refrigerant compressor in which an increased load torque acts on the compression mechanism (compression phase) and the at least one transit angle segment is provided in each operating phase of the refrigerant compressor in which a reduced load torque with respect to the operating phase with increased load torque acts on the compression mechanism (intake phase).

In cases in which the load torque varies particularly strongly due to the pressure differences prevailing in the refrigerant circuit during a crankshaft rotation, it is provided according to the invention that the compression mechanism is subjected to no positive operating torque during the at least one transit angle segment instead of the reduced positive operating torque. In this case, it is even possible to apply a braking torque instead of the positive operating torque on the compression mechanism.

Overall, this produces the possibility to equalize the non-uniform profile of the rotational frequency and the load torque with respect to the crank angle in a targeted way. In addition, there is also the possibility to selectively feed a (higher) positive operating torque to the compression mechanism when this is necessary to overcome the resulting load torque. Especially at low rotational speeds, the recurring shocks transferred by the load torque to the compression mechanism can be selectively damped, and even completely compensated for.

This arrangement not only enables operation at low rotational speeds without the risk of exciting the oscillation system in the natural frequency range, but a reversal of the direction of rotation of the compression mechanism can also be effectively prevented without having to increase its rotational speed overall.

The duration of the regulating time interval is tailored to the requirements on the uniformity of the rotational movement of the compression mechanism and can correspond either to the total operating period of the refrigerant compressor or else to only part of this period. In an especially preferred way, the regulating time interval begins when a signal of the electronic control device of a refrigerator signals that a target temperature in a cooling compartment has been reached and ends at the same time as the drive unit is switched off (switch-off time).

In this case, the electronic control device is also designed to reduce the rotational speed of the drive unit and thus the compression mechanism during the regulating time interval, even though a positive operating torque is applied on the compression mechanism depending on the crankshaft rotation over a defined crank angle range. In addition to making the rotational frequency uniform with respect to the crank angle, a constant reduction of the rotational speed of the compression mechanism can also be achieved in this way.

By means of the electronic control device according to the invention, it is also possible to increase the rotational speed of the drive unit, so that a constant acceleration of the compression mechanism can also be achieved in this way, which is advantageous, in particular, for the starting phase.

According to one especially preferred variant of the invention, the ratio between positive operating torque and reduced positive operating torque ≥1/0.2, preferably ≥1/0.1, and especially preferably ≥1/0.03. Because the determination of the rotational speed of a refrigerant compressor with variable rotational speed, and thus the regulation, are performed in an especially convenient way based on the countervoltage induced in the motor winding, it is advantageous to drive the refrigerant compressor in the at least one transit angle segment with positive operating torque, even if it is very small, compared to the operating torque in the drive angle segment, because in this way an especially stable detection of the induced countervoltage is possible. In the normal case, it is sufficient that the ratio between the positive operating torque and the reduced positive operating torque is greater than 1/0.03, i.e., when the reduced positive operating torque is less than 3% of the positive operating torque in the drive time segment [sic; drive angle segment], in order to still be able to detect a sufficiently stable induced countervoltage. At the same time, this enables a low operating torque so that the rotational frequency of the compression mechanism practically does not increase during the transit angle segment, which enables the use of the electronic control device according to the invention, especially for the uniform reduction of the rotational speed before the drive is switched off.

According to another preferred embodiment of the invention, it is provided that the electronic control device is designed to alternately provide multiple drive angle segments and multiple transit angle segments during one crankshaft rotation. In this way, it is possible to be responsive in an especially specific way to the load torque varying with respect to one crankshaft rotation. The individual drive and transit angle segments can have different sizes here, thus they can cover different crank angles and can be completely tuned together with the ratio between positive operating torque and reduced positive operating torque to the profile of the load torque, so that non-uniformity of the rotational frequency with respect to the crank angle can be compensated not only approximately but nearly completely, and the compression mechanism also rotates very uniformly at low rotational speeds. In other words, due to the electronic control device according to the invention for reciprocating piston refrigerant compressors, a very harmonic rotational movement of the compression mechanism can be achieved, because the rotational frequency can be kept constant or approximately constant with respect to the crank angle, but in any event, sufficiently constant that excitation of the oscillation system in the natural frequency range can be avoided.

According to another preferred variant of the invention, it is provided that the electronic control device is designed to provide exactly one drive angle segment and exactly one transit angle segment during one complete crankshaft rotation, wherein, in this case, preferably the drive angle section and transit angle section together comprise one crankshaft rotation, that is, 360°. This is an especially simple method of control, in which, during one crankshaft rotation, the operating torque of the drive unit is increased exactly once and is decreased again, preferably to the value described above, or else is completely eliminated during the rest of the crankshaft rotation.

In one especially preferred variant of the invention, it is provided that the exactly one drive angle segment is provided during a crank angle from 220° to 360°, especially preferably during a crank angle from 270° to 360°. In this range, the piston is in the compression phase, and feeding the positive operating torque in this phase during a crank angle between 220° and 360° is sufficient in most cases in practice to sufficiently damp the shocks acting on the compression mechanism due to the load torque for low rotational speeds of the compression mechanism, in order to prevent strong deflections of the spring element. In addition, the feeding of a positive operating torque in this segment of the crank angle can also be used to protect the piston from a reversal of the direction of rotation, if the kinetic energy of the unit consisting of the compression mechanism and drive unit in the driveless state is no longer sufficient to overcome the load torque.

In an especially preferred way, the electronic control device is designed to provide the drive angle segment or segments with positive operating torque when the piston is located basically between a lower and an upper dead point in a compression phase, because in this way, the piston can be driven very precisely according to the load torque acting on it. Even if the rotational speed of the compression mechanism should slow overall, in order to be subsequently switched off, due to the provision of the drive angle segment or segments during the compression phase, it is possible to drive the piston so that it can just end the compression phase without the risk of the piston striking backward due to the lack of sufficient kinetic energy.

Furthermore, the electronic control device can then provide the transit angle segments with reduced positive operating torque or without any positive operating torque, when the piston is located in a suction phase between an upper and a lower dead point. The compression mechanism therefore experiences no or only very slight drive by the drive unit in the intake phase.

Thus, overall, not only can a uniform rotation of the compression mechanism be realized with constant or nearly constant rotational frequency during one crankshaft rotation, and thus also damping of the shocks caused by the load torque in the compression phase even at low speeds, but it is also possible to shift the switch-off time to a rotational speed that is very low, for example, 250 rpm, because passing through the low rotational speeds during the stopping process no longer ca elated problems for the reasons specified above.

According to another preferred variant of the invention, it is provided that the electronic control device is designed to switch the drive unit to an unpowered state at a switch-off time, so that this no longer generates positive operating torque, in order to let the compression mechanism run out to a standstill.

In an especially preferred way, the electronic control device is designed such that it provides the regulating time interval immediately before the switch-off time. In this way, the rotational speed of the compression mechanism can be regulated down to a very low rotational speed with the advantages described above and the refrigerant compressor can then be switched off, the electronic control device is preferably designed to select the switch-off time so that the kinetic energy of the unit consisting of the compression mechanism and the drive unit is sufficient at the switch-off time to enable the piston to overcome at least the next upper dead point following the switch-off time and/or so that the piston of the compression mechanism comes to a standstill after reaching the next upper dead point following the switch-off time and before reaching another lower dead point immediately following this upper dead point and/or so that the piston of the compression mechanism comes to a standstill after the next upper dead point following the switch-off time and before reaching a crank angle of 220° following this next upper dead point. In this way, it is guaranteed that the piston of the compression mechanism can nevertheless run out in an unregulated and unbraked way despite the circumstance that no braking torque is applied to the compression mechanism, without there being the risk of the piston striking backward or, even if the piston does strike back, the impulse thus exerted on the compression mechanism does not cause an impermissibly large deflection of the spring element.

According to another preferred variant of the invention, it is provided that the electronic control device is designed to provide the positive operating torque of each drive time segment [sic; drive angle segment] and the reduced positive operating torque relative to the positive operating torque of each transit time segment [sic; transit angle segment] for each crank angle of one crankshaft rotation of the regulating time interval such that it corresponds to the load torque 1 m acting on the compression mechanism at each crank angle. In this way, it is possible to, so to speak, mirror the load torque and exert it as an operating torque on the compression mechanism, whereby an especially uniform rotational frequency can be achieved with respect to the crank angle.

The goals mentioned above are also achieved by means of a method for regulating a reciprocating piston refrigerant compressor, whose compression mechanism is driven by means of a drive unit with an operating torque, in that the method has the following steps

during a regulating time interval (Δt) having more than one crankshaft rotation

for multiple, preferably for all crankshaft rotations of the regulating time interval (Δt), the following is performed each time

detect the position of the crankshaft (6) or the piston (9) of the reciprocating piston refrigerant compressor (1)

compare the detected position with at least one prespecified reference position

starting from the at least one prespecified reference position, drive the compression mechanism (5) with a positive operating torque (Bm) for the duration of at least one drive angle segment (ΔΦ) of one crankshaft, rotation

drive the compression mechanism (5) with a reduced positive operating torque (Bmv) relative to the positive operating torque (Bm) or no positive operating torque for the duration of at least one transit angle segment (Δτ).

According to one preferred variant of the method according to the invention, it is provided that the ratio between positive operating torque (Bm) and reduced positive operating torque (Bmv) is ≥1/0.2, preferably ≥1/0.1, and especially preferably ≥1/0.03.

According to another preferred variant of the method according to the invention, it is provided that multiple reference positions are provided.

According to another preferred variant of the method according to the invention, it is provided that exactly one drive angle segment and one transit angle segment are provided during one crankshaft rotation.

According to another preferred variant of the method according to the invention, it is provided that the at least one reference position is provided during a crank angle from 220° to 360°, especially preferably during a crank angle from 270° to 360°.

According to another preferred variant of the method according to the invention, it is provided that the rotational speed of the reciprocating piston refrigerant compressor is reduced or increased during the regulating time interval.

According to another preferred variant of the method according to the invention, it is provided that the regulating time interval is provided immediately before a switch-off time, after which the drive unit is switched to an unpowered state, preferably starting the regulating time interval when a signal of the electronic control device of a refrigerator signals that a target temperature has been reached in a cooling compartment.

BRIEF DESCRIPTION OF THE FIGURES

The invention will now be explained in more detail using one or more embodiments with reference to the following figures.

FIG. 1, a schematic representation of a reciprocating piston refrigerant compressor in a refrigerant circuit,

FIG. 2, a schematic view of a compression mechanism,

FIG. 3, a diagram relating to the load torque profile and operating torque profile with respect to the crank angle in a reciprocating piston refrigerant compressor according to the prior art,

FIG. 4, a diagram relating to the angular acceleration of the crankshaft with respect to the crank angle in a reciprocating piston refrigerant compressor according to the prior art,

FIG. 5, a diagram relating to the load torque profile and operating torque profile with respect to the crank angle using an electronic control device according to the invention,

FIG. 6, a diagram relating to the angular acceleration of the crankshaft with respect to the crank angle using an electronic control device according to the invention,

FIG. 7, a diagram relating to the load torque profile and operating torque profile with respect to the crank angle with mirrored operating torque,

FIG. 8, a diagram relating ,h- angular acceleration of the crankshaft with respect to the crank angle with mirrored operating torque,

FIG. 9, a diagram relating to the load torque profile and operating torque profile with respect to the crank angle during a stopping process, and

FIG. 10, a diagram relating to the load torque profile and operating torque profile with respect to time during a stopping process.

FIG. 1 shows a schematic diagram of a reciprocating piston refrigerant compressor 1 connected to an electrical power supply 12 and regulated by means of an electronic control device 13 in a known coolant circuit with a condenser 2 of a throttle device 3 and an evaporator 4. The refrigerant removes heat from the cooling compartment in the evaporator 4, which cools the compartment. The evaporated refrigerant is compressed, by means of the compression mechanism 5 of the reciprocating piston refrigerant compressor 1, to a higher temperature, and consequently becomes a liquid again in the condenser 2, in order to be finally fed via the throttle device 3 back to the evaporator 4 of the cooling compartment.

In the present embodiment, the electronic control device 13 of the refrigerant compressor I communicates with an electronic control device 14 of a refrigerator 15. However, such a communications ability is not considered essential to the invention, because it is also conceivable that the electronic control device 13 communicates with a refrigerator 15 that itself does not have its own electronic control device, but instead only a thermostat.

FIG. 2 shows a schematic view of the compression mechanism 5 consisting of a crankshaft 6 driven by means of a drive unit 18, a connecting rod 7, and also a piston 9 that can move back and forth in a cylinder block 8. The compression mechanism 5 is supported by means of spring element 10 in a housing 11, wherein this spring element 10 is intended to absorb and compensate for oscillations occurring in the unit consisting of compression mechanism 5 and drive unit 18 due to the rotation of the crankshaft 6 and movements of the piston 9.

The drive unit 18 control led by the electronic control device 13 is a drive unit 18 with variable rotational speed, typically a brushless direct-current motor, whose rotational speed can be regulated by means of the electronic control device 13. The detection of the instantaneous rotational speed required for regulating the rotational speed is performed by detecting the countervoltage (induction countervoltage) induced in the motor winding, so that no other sensors are required, even though the electronic control device 13 according to the invention obviously can also interact with separate sensors for measuring the rotational speed, for example, Hall sensors.

During the operating period of a reciprocating piston refrigerant compressor with variable rotational speed, a distinction is to be made between basically 3 phases:

the starting phase

the normal regulated operating phase

the stopping process.

The basic condition is a cooling compartment temperature (=target temperature) of the refrigerator 15 that can be preselected by the user of a refrigerator 15 within limits. If a cooling compartment cooled to the target temperature is assumed and a load is placed in the refrigerator 15 or the refrigerator door is opened, warm air flows into the cooling compartment. The electronic control device 14 of the refrigerator 15 detects that the cooling compartment temperature is increased and sends a signal (usually a frequency signal) to the electronic control device 13 of the refrigerant compressor 1, with which this control device is informed that cooling power is needed, whereupon this control device controls and regulates the refrigerant compressor 1 in accordance with its programming, in order to supply (more or less) cooling power.

In the example concerning an object, the electronic control device 13 of the refrigerant compressor 1 starts this compressor order to compress the refrigerant and remove heat from the cooling compartment and in order to restore the target temperature. This “startup” initiates the starting phase. Here, the refrigerant compressor 1, more specifically its drive unit 18, is accelerated to a certain rotational speed prespecified by the electronic control device 13 of the refrigerant compressor 1. Reaching this rotational speed ends the starting phase. At this time, the target temperature has usually not yet been reached.

The refrigerant compressor then transitions into the normal regulated operating phase. This continues as long as the refrigerant compressor 1 is switched on or, expressed somewhat more technically, as long as energy is supplied to the refrigerant via the compression mechanism 5 and the drive unit 18 of the refrigerant compressor 1 generates an operating torque. The compression mechanism 5 can rotate during this normal regulated operating phase at a different rotational speed, according to whether more or less heat is to be removed from the cooling compartment. For example, if someone opens the doors of the refrigerator 15 during such a normal regulated operating phase, due to the incoming warm air, the electronic control device 14 of the refrigerator 15 demands more cooling power from the refrigerant compressor 1, so that the electronic control device 13 of the refrigerant compressor 1 increases the rotational speed of the drive unit 18, and thus of the compression mechanism 5, in order to be able to dissipate the heat flowing into the cooling compartment.

Increasing the rotational speed is associated with increased energy demands on the refrigerant compressor 1. If the electronic control device 14 of the refrigerator 15 detects that the current cooling compartment temperature is approaching the target temperature, the electronic control device 14 of the refrigerator 15 will send a corresponding signal to the electronic control device 13 of the refrigerant compressor 1, in order to request less cooling power and to not “shoot past” the target temperature and to approach this target temperature slowly. The electronic control device 13 of the refrigerant compressor 1 will, in turn, reduce the rotational speed of the drive unit 18 of the compression mechanism 5 due to this request.

If the electronic control device 14 of the refrigerator 15 detects that, in the meantime, the cooling compartment temperature has increased again, because, for example, new loads were placed in the cooling compartment, then the electronic control device H of the refrigerator 15 will again request more cooling power from the electronic control device 13 of the refrigerant compressor 1, so that this will again increase the rotational speed of the drive unit 18 of the compression mechanism 5.

If, after an appropriate duration of the normal regulated operating phase, the target temperature is reached, the electronic control device 14 of the refrigerator 15 sends a signal to the electronic control device 13 of the refrigerant compressor 1, with which this control device is informed that the target temperature has been reached. Then the electronic control device 13 of the refrigerant compressor 1 switches off the drive unit 18 (switch-off time AZ). Switching off the drive unit 18 has the result that the compression mechanism 5 is located together with the drive unit 18 in a driveless state, and continues to rotate only due to mass inertia until the rotational speed is 0. Colloquially, it could also be said that the refrigerant compressor “runs out.”

During the operation of the compression mechanism, shocks can be generated that are exerted by the load torque during the compression phase on the compression mechanism 5 and that repeat with each crankshaft rotation and can coincide at low rotational speeds with the natural frequency of the oscillation system formed by the spring element 10, whereby the deflection of this spring element increases such that it can lead to contact between the unit consisting of compression mechanism 5 and drive unit 18 with the housing 11, whereby undesired noise emissions are generated.

In addition, during the stopping process, when the drive unit 18 no longer generates operating torque, a reversal of the direction of rotation of the compression mechanism 5 can occur, whereby an additional shock is exerted on the compression mechanism that likewise results in an undesirably strong deflection of the spring element 11 [sic; 10], with the consequence that this reversal of the direction of rotation also produces the risk that the unit consisting of the compression mechanism 5 and drive unit 18 come into contact with the housing 11 and produce sound emissions.

In summary, it can be stated that low rotational speeds, independent of whether the refrigerant compressor is in the starting phase, the normal regulated operating phase, or the stopping process, always produce the risk that the oscillation system formed by the spring element 10 will be excited in the range of its natural frequency and will therefore produce the described contacts that produce noise between the unit consisting of the compression mechanism 5 and drive unit 18 and housing 11.

FIG. 3 shows a diagram of the profile of the load torque Lm with respect to the crank angle Φ during a normal regulated operating phase of a reciprocating piston refrigerant compressor known from the prior art, whose drive unit 18 drives the compression mechanism 5 with an operating torque Bm. Here, it was assumed that the crankshaft rotates in the clockwise direction. The direction of rotation thus goes from 0° (upper dead point) to 360° (upper dead point).

As shown from the diagram, the load torque Lm is greatest shortly before the piston 9 reaches the upper dead point in the compression phase, that is, at approx. 330°, and is negative at the beginning of the intake phase, that is, in the present case, at approx. 10°, i.e., the load Lm in this section of the intake phase (re-expansion phase) supports the rotation of the compression mechanism 5.

FIG. 4 shows a diagram in which the angular acceleration {acute over (ω)} of the crankshaft 6 with respect to the crank angle Φ is plotted as a result of the ratio between the load torque Lm and operating torque Bm as shown in FIG. 3.

It can be seen that the maximum negative angular acceleration {acute over (ω)} of the crankshaft 6 occurs at the time of the maximum load torque Lm, while the angular acceleration {acute over (ω)} is positive during the intake phase and at the beginning of the compression phase up to approx. 250°, so that the rotational speed of the crankshaft 6 is increased in this crank angle range.

The angular acceleration {acute over (ω)} here varies in the present embodiment between the values of approx. −3400 rad/s² and approx. +1000 rad/s². This circumstance has the result that, despite the applied operating torque Bm with respect to one crankshaft rotation, the rotation of the compression mechanism 5, especially the crankshaft 6, is very uneven, and in the range of a crank angle of approx. 330° the load torque Lm exerts a shock on the compression mechanism 5, wherein this shock repeats for every crankshaft rotation and the oscillations already described in detail above are generated, with the similarly already described negative effects, wherein the load torque Lm increases with decreasing rotational speed ω.

In practice, reciprocating piston refrigerant compressors are therefore not operated at rotational speeds that are in the range of the natural frequencies of the oscillation system.

To nevertheless enable operation even at low rotational speeds without the risk of disruptive noise emissions, it is provided according to the invention that the electronic control device 13 is designed so that it actively varies the operating torque Bin during a regulating time interval Δt at least once per crankshaft rotation as a function of the crank angle, by increasing the voltage supply of the drive unit 18 during a drive angle segment ΔΦ relative to the rest of the crankshaft rotation.

FIG. 5 shows the profile of the load torque Lm known from FIG. 3 during one crankshaft rotation, but with schematically shown operating torque Bm that is applied according to the invention by the electronic control device 13 and is regulated so that exactly one drive angle segment ΔΦ is provided, during which the drive unit 18 drives the compression mechanism 5 with a positive operating torque Bm, and exactly one transit angle segment Δτ, during which the drive unit 18 drives the compression mechanism 5 with a reduced positive operating torque Bmv relative to the positive operating torque or, alternatively, does not drive at all (see dashed line of Bmv), wherein the transitions between the positive operating torque Bm and the reduced operating torque Bmv are indeed drawn, but are not provided with an extra reference symbol, but instead are allocated to the positive operating torque Bm for the sake of simplicity.

The ratio between the positive operating torque Bm and reduced positive operating torque Bmv is preferably −1/0.03 in this embodiment. Typical pressure ratios in reciprocating piston refrigerant compressors, however, also permit ratios from ≥1/0.1 or ≥1/0.2 without deviating from the concept of the invention.

In FIG. 5, the beginning of the compression phase KP is also marked with the reference symbol 16 and the beginning of the intake phase is marked with the reference symbol 17. As can also be seen from FIG. 5, the electronic control device 13 of the refrigerant compressor 1 regulates the drive unit 18 so that the drive angle segment ΔΦ is completely or for the most part within the compression phase. Depending on the (average) magnitude of the positive operating torque Bm and thus on the output capability of the drive unit 18, the drive angle segment ΔΦ can be greater or less than shown in the embodiment.

As can be seen from FIG. 6, which shows a diagram relating to the angular acceleration {acute over (ω)} of the crankshaft 6 with respect to the crank angle Φ for the use of an electronic control device 13 according to the invention with an operating torque profile according to FIG. 5, this driving of the compression mechanism 5, both applied over a defined crank angle and also controlled with respect to magnitude during one crankshaft rotation with two different operating torques Bm and Bmv, leads to the result that the angular acceleration {acute over (ω)} of the crankshaft 6 with respect to the crank angle Φ (and thus also the rotational frequency with respect to the crank angle) has a significantly more even profile than is the case for conventional reciprocating piston refrigerant compressors according to the prior art and shown in FIG. 4. In the present embodiment, the rotational frequency in the crank angle range between 30° and 210° is approximately constant, so the crankshaft experiences no acceleration. In the range of the crank angle between 210° and 360°, the crankshaft is initially accelerated and then braked again, so that the evenness of the rotational frequency of the crankshaft is influenced only insignificantly and thus also the shocks acting on the compression mechanism are significantly lower than was previously the case.

In another variant of the invention, it can be provided that the profile of the positive operating torque Bm and the reduced positive operating torque Bmv together, with which the compression mechanism 5 is driven, corresponds practically to the profile of the load torque Lm, to which the compression mechanism 5 is exposed, but with the opposite sign and possibly with larger or smaller magnitudes in some sections, according to whether the rotational speed of the compression mechanism should be kept constant, increased, or reduced.

FIG. 7 shows a matched combination of positive operating torque Bm and reduced positive operating torque Bmv. As can be seen from the diagram, the positive operating torque Bm varies both in the drive angle segment ΔΦ and also in the transit angle segment Δτ, so that the profile of the load torque Lm is reproduced, but with the opposite sign.

FIG. 8 shows that, in this case, the angular acceleration {acute over (ω)} of the crankshaft 6 is constant, whereby a constant rotational frequency is produced with respect to the crank angle Φ and thus a constant rotational speed is also produced.

Another advantage of the electronic control device 13 according to the invention is the ability to optimize the stopping process of a reciprocating piston refrigerant compressor 1, in that this compressor

can be either actively run down until a full standstill, without having to actively brake the compression mechanism 5 and without causing a rotation of the compression mechanism 5 in the opposite direction or

can be actively run down until just before standstill, but in any even until rotational frequencies are below 450 rpm, preferably below 250 rpm, where the drive unit 18 can be itched off without a problem, because at these low rotational frequencies, the piston striking backward, if it does so at all, no longer has any negative noise-related effects.

Here, the profile of the positive operating torque Bm and the reduced positive operating torque Bmv is not matched to the profile of the load torque Lm such that the first profile is practically equivalent to a mirrored load torque Lm (see FIGS. 7 and 8), because this would lead to a constant rotational frequency and constant rotational speed overall.

During the stopping process, however, the rotational speed should be slowed independent of the fluctuation of the rotational frequency with respect to the crank angle Φ, so that the electronic control device 13 exerts the operating torque lam and the reduced positive operating torque Bmv so that the shocks occurring from the load torque Lm during the compression phase on the compression mechanism are dissipated for the most part, but the rotational speed overall decreases.

Simultaneously, the positive operating torque Bm can be selected so that, in any event, in the compression phase, there is always sufficient drive to prevent the piston 9 from striking backward.

Here it can also be provided, in contrast to the shown embodiments, that the drive angle segment ΔΦ begins in the intake phase of the piston 9, so that the piston 9 is already provided with sufficient momentum before entry in the compression phase, in order to overcome the subsequent upper dead point, without causing backward striking by the load torque Lm.

Through appropriate magnitude matching of the positive operating torque Bm and the reduced positive operating torque Bmv to each other, the compression mechanism 5 can slow its rotational movement overall up to full standstill or up to a time at which a possible backward striking of the piston 9 would no longer have any negative noise-related effects.

Here, it is essential that the electronic control device 13 can also, in the course of reducing the rotational speed for the purpose of preparing the switching off of the drive unit 18, provide at least one such drive angle segment ΔΦ and at least one such transit angle segment Δτ at least during a regulating time interval Δt lasting multiple crankshaft rotations.

FIG. 9 shows a diagram relating to the profile of the load torque Lm, the positive operating torque Bm, the reduced positive operating torque Bmv, the rotational speed ω, and the angular acceleration {acute over (ω)} of the compression mechanism 5 beyond multiple crankshaft rotations and over a regulating time interval Δt, wherein the rotational speed ω is continuously reduced until the drive is switched off at the switch-off time AZ at a rotational speed below 450 rpm, preferably below 250 rpm. The switching off is here performed preferably during the compression phase.

According to the invention it is provided that the electronic control device 13 continuously drives the compression mechanism with a positive operating torque Bm during a drive angle segment ΔΦ and a reduced positive operating torque Bmv during a transit angle segment Δτ, instead of switching off the drive unit 18 realized in the prior art at relatively high rotational speeds and subsequently letting it run out, with the optional additional application of a braking torque, wherein the positive operating torque Bm preferably remains constant, optionally slightly increases, in terms of magnitude for each crankshaft rotation up to the switch-off time AZ, in order to compensate for the loss of speed and thus the reduction in the kinetic energy.

Preferably, the positive operating torque Bm remains constant, but is applied, in any event, with decreasing rotational speed over a longer time span, thus, also during an increasing crank angle.

In this way, the compression mechanism is always compressed somewhat more or somewhat longer or both, in order to be able to overcome the load torque and to compensate for the “lost” kinetic energy of the unit consisting of the compression mechanism 5 and drive unit 18. Switching off the drive unit 18 takes place only at a very low rotational speed below 450 rpm, preferably below 250 rpm.

The regulating time interval Δt ends in the present embodiment at the switch-off time AZ or, in other words, the regulating time interval Δt immediately precedes the switch-off time AZ.

As can be seen from FIG. 9, after the switch-off time AZ, the piston 9 strikes backward, because the compression mechanism 5 is no longer exposed to operating torque during the stopping process and the piston 9 can no longer overcome the load torque Lm in the compression phase. Consequently, a pendulum motion of the crankshaft 6 about the lower dead point is realized, wherein the occurring load torques Lm are small in terms of magnitude such that these no longer cause the spring element 10 to deflect, which leads to contact between the unit consisting of the compression mechanism 5 and drive unit 18 on one hand and the housing 11 on the other hand.

FIG. 10 corresponds to FIG. 9, but with the difference that the abscissa is not the crank angle Φ, but instead the time t. It can be seen that with increasing rotational speed ω, the time period over which the positive operating torque Bm is applied, increases.

The rotational speed reaches an inflection point at approx. t=0.78, which corresponds to the backward striking of the piston 9. The pendulum motion visible from FIG. 9 about the lower dead point and the associated continuous changing of the rotational speed to between a positive value and a negative value is no longer seen in FIG. 10.

LIST OF REFERENCE SYMBOLS

1 Reciprocating piston refrigerant compressor

2 Condenser

3 Throttle device

4 Evaporator

5 Compression mechanism

6 Crankshaft

7 Connecting rod

8 Cylinder block

9 Piston

10 Spring element

11 Housing

12 Power supply

13 Electronic control device of the refrigerant compressor

14 Electronic control device of the cooling chamber

15 Cooling chamber

16 Start of the compression phase

17 Start of the intake phase

18 Drive unit 

1. An electronic control device for a refrigerant compressor, comprising at least a drive unit, a compression mechanism that is actively connected to the drive unit, with at least one piston that is driven by a crankshaft and moves back and forth between a lower and an upper dead point in a cylinder of a cylinder block, wherein the electronic control device is designed to detect and to control and/or regulate the rotational speed (ω) of the drive unit, and to detect the piston position at least approximately, wherein the electronic control device as designed to drive the compression mechanism by means of the drive unit such that, for the duration of one regulating time interval (Δt) having more than one crankshaft rotation, for multiple crankshaft rotations, at least one drive angle segment (ΔΦ) and at least one transit angle segment (Δτ) are provided, and wherein the compression mechanism is exposed during the at least one drive angle segment (ΔΦ) to a positive operating torque (Bm), and during the at least one transit angle segment (Δτ) to a reduced positive operating torque (Bmv) relative to the positive operating torque (Bm), or to no positive operating torque (Bm).
 2. The electronic control device of a refrigerant compressor according to claim 1, wherein the ratio between positive operating torque (Bm) and reduced positive operating torque (Bmv) is ≥1/0.2.
 3. The electronic control device according to claim 1, wherein it is designed to provide multiple drive angle segments (ΔΦ_(n)) and multiple transit angle segments (Δτ_(n)) alternately during one crankshaft rotation.
 4. The electronic control device according to claim 1, wherein it is designed to provide the drive angle segment or segments (ΔΦ, ΔΦ_(n)) with positive operating torque (Bm) when the piston is located between a lower and an upper dead point in a compression phase.
 5. The electronic control device according to claim 4, wherein it is designed to provide exactly one drive angle segment (ΔΦ) and one transit angle segment (Δτ) during one complete crankshaft rotation.
 6. The electronic control device according to claim 1, wherein it is designed to provide the transit angle segment or segments (Δτ, Δτ_(n)) with reduced positive operating torque (Bmv) or with no positive operating torque when the piston is located between an upper and a lower dead point in an intake phase.
 7. The electronic control device according to claim 1, wherein it is designed to apply a braking torque to the compression mechanism during the transit angle segment or segments (Δτ, Δτ_(n)).
 8. The electronic control device according to claim 1, wherein it is designed to reduce or increase the rotational speed (ω) of the drive unit during the regulating time interval (Δt).
 9. The electronic control device according to claim 1, wherein it is designed to switch the drive unit to a powerless state at a switch-off time (AZ), so that this drive unit no longer generates positive operating torque (Bm), in order to let the compression mechanism run out to a standstill.
 10. The electronic control device according to claim 9, wherein it is designed to provide the regulating time interval (Δt) immediately before the switch-off time (AZ).
 11. The electronic control device according to claim 9, wherein it is designed to select the switch-off time (AZ) so that the kinetic energy of the compression mechanism at the switch-off time (AZ) is sufficient to enable the piston to overcome at least the next upper dead point following the switch-off time (AZ).
 12. The electronic control device according to claim 9, wherein it is designed to select the switch-off time (AZ) so that the piston of the compression mechanism comes to a standstill after reaching the next upper dead point following the switch-off time (AZ) and before reaching the lower dead point immediately following this upper dead point.
 13. The electronic control device according to claim 9, wherein it is designed to select the switch-off time (AZ) so that the piston of the compression mechanism comes to a standstill after the next upper dead point following the switch-off time (AZ) and before reaching a crank angle of 220° following this next upper dead point.
 14. The electronic control device according to claim 1, wherein it is designed to form the positive operating torque (Bm) during the at least one drive angle segment (ΔΦ) and/or the reduced positive operating torque (Bmv) during the at least one transit angle segment (Δτ), each with varying magnitude.
 15. The electronic control device according to claim 14, wherein it is designed to form the common profile of the positive operating torque (Bm) and the reduced positive operating torque (Bmv) for each crankshaft rotation so that it corresponds to the load torque (Lm) acting on the compression mechanism during this crankshaft rotation.
 16. A hermetically encapsulated refrigerant compressor with an electronic control device according to claim
 1. 17. The hermetically encapsulated refrigerant compressor according to claim 16, wherein an acceleration sensor is provided on the drive unit and/or a pressure sensor is provided in the cylinder of the cylinder block.
 18. A method for regulating a reciprocating piston refrigerant compressor, whose compression mechanism is driven with an operating torque by means of a drive unit, comprising the following steps during a regulating time interval (Δt) having more than one crankshaft rotation; for multiple crankshaft rotations of the regulating time interval (Δt), the following is performed each time; detect the position of the crankshaft or the piston of the reciprocating piston refrigerant compressor; compare the detected position to at least one prespecified reference position; starting from the at least one prespecified reference position, drive the compression mechanism with a positive operating torque (Bm) for the duration of at least one drive angle segment (ΔΦ) of one crankshaft rotation; drive the compression mechanism with a reduced positive operating torque (Bmv) relative to the positive operating torque (Bm) or no positive operating torque for the duration of at least one transit angle segment (Δτ).
 19. The method according to claim 18, wherein the ratio between positive operating torque (Bm) and reduced positive operating torque (Bmv) is ≥1/0.2.
 20. The method according to claim 18, wherein multiple reference positions are provided during one crankshaft rotation.
 21. The method according to claim 18, wherein exactly one drive angle segment (ΔΦ) and one transit angle segment (Δτ) are provided during one crankshaft rotation.
 22. The method according to claim 18, wherein the at least one reference position is provided during a crank angle from 220° to 360°.
 23. The method according to claim 18, wherein the rotational speed of the reciprocating piston refrigerant compressor is reduced or increased during the regulating time interval (Δt).
 24. The method according to claim 18, wherein the regulating time interval (Δt) is provided immediately before a switch-off time (AZ), after which the drive unit is switched to a powerless state.
 25. The electronic control device of claim 1, wherein the multiple crankshaft rotations comprise each crankshaft rotation of the regulating time interval (Δt).
 26. The electronic control device of claim 2, wherein the reduced operating torque is ≥1/0.1.
 27. The electronic control device of claim 2, wherein the reduced operating torque is ≥1/0.03.
 28. The electronic control device of claim 5, wherein the exactly one drive angle segment (ΔΦ) is provided during a crank angle from 220° to 360°.
 29. The electronic control device of claim 5, wherein the exactly one drive angle segment (ΔΦ) is provided during a crank angle from 270° to 360°.
 30. The electronic control device of claim 10, wherein the regulating time interval (Δt) is begun when a signal of the electronic control device of a refrigerator signals that a target temperature has been reached in a cooling compartment.
 31. The method of claim 18, wherein the multiple crankshaft rotations comprise all crankshaft rotations.
 32. The method of claim 19, wherein the reduced operating torque (Bmv) is ≥1/0.1.
 33. The method of claim 19, wherein the reduced operating torque (Bmv) is ≥1/0.03.
 34. The method of claim 22, wherein the crank angle comprises 270° to 360°.
 35. The method of claim 24, wherein the regulating time interval (Δt) begins when a signal of the electronic control device of a refrigerator signals that a target temperature has been reached in the cooling compartment. 